Percussive unit for earth drilling



March 7, 1967 J. M,.\fcLEARfY PERCUSSIVE UNIT` FOR EARTH DRILLING Filed July 28, 1966 5 Sheets-Sheet 1 wIH l INVENTOR.

vJames M. Cleary Attorney March 7;. I

J; M: @LEAN-` PERCUSSIVE UNIT4 FOR EARTH DRILLING 5 Sheets-Sheet 2 Filed July 28, 1966 /NVE/VTOH James M. Cleary .5y @www/law Attorney United States Patent() 3,307,640 PERCUSSIVE UNIT FOR EARTH DRILLING James M. Cleary, Dallas, Tex., assgnor to Atlantic Richfield Company, Philadelphia, Pa., a corporation of Pennsylvania Filed July 28, 1966, Ser. No. 568,436 36 Claims. (Cl. 173-137) This is a continuation-in-part of application Serial` No. 511,870, tiled December 6, 1965, now Patent No. 3,270,822, which is a continuation-in-part of application Serial No. 241,412, led November 30, 1962, now Patent No. 3,229,775.

This invention is concerned with fluid-actuated, percussive units for earth borehole drilling. More specilically, this invention is concerned with a ported rotatable tubular valve whose exterior surface between the ports is elliptically shaped and with a percussive unit employing this valve.

Fluid-actuated earth drilling percussive units are commonly composed of an elongated tubular housing having a power fluid inlet and outlet, a valving and passage system for controlling passage of the power fluid through the power unit, and a power chamber. Slidably mounted in the power chamber is a piston-like hammer which is adapted to reciprocate up and down to impact an anvil to which a drill bit is attached. Longitudinal reciprocal movement of the hammer is accomplished by alternately applying power fluid to the ends of the hammer. When power uid is applied to the lower end of the hammer between the anvil and hammer, the hammer is accelerated in an upward direction until entry of the power fluid no longer acts on the lower end ofthe hammer. The upward energy of lthe hammer is then transmitted to an elastic rebound cushion. This rebound cushion is usually either a trapped volume of compressible power fluid, or a hammer return spring, or both. The hammer continues moving upward until the velocity of the hammer is zero and the direction of movement of the hammer is reversed. Before or at the peak of the upstroke of the hammer, the valving system switches flow of high pressure power fluid to the upper end of the hammer. When the hammer starts to move downward, part of the energy stored in the elastic rebound cushion is returned to the hammer. This energy combines with energy developed by the power fluid to accelerate the hammer downward to impact the anvil. This downward percussive force is transmitted to the drill bit.

The ultimate objective in percussive tool design is to provide a durable gasor liquid-operated tool exhibiting both a high energy, high frequency impact and a high fluid conductivity per unit diameter over a wide range of operating conditions. The achievement of this ultimate objective depends primarily on the stability, reliability and uniformity of the timing relationship between the up and down movement of the hammer and the instant that the valving system ceases and starts the flow of power fluid to alternate ends of the hammer.

Patent No. 3,229,775 provided a novel valving system employing a ported tubular rotatable valve for controlling passage of power fluid to and from the opposite ends of the hammer Copending application Serial No. 511,870 combined the ported, rotatable tubular valve with motion conversion means, e.g. a cam and follower, adapted to employ the reciprocating or longitudinal movement of a hammer to coordinate the axial motion and position of the hammer with the circumferential rotative position of the ports in the valve during at least a portion of the reciprocating cycle of the hammer. The preferred rotatable valve has at least two inlet ports and at least two exhaust ports which ports are spaced substantially symmetrically around the circumference of the valve.

ICC

The valve may also be connected to a rotary drive means which rotates the valve at a speed of rotation greater than the speed' of rotation of the percussive unit and drill bit. The timing relationship between the hammer and rotatable tubular valve is further improved by providing a rebound spring above the hammer. Preferably, this spring is a rod-like spring.

In the preferred arrangements of Copending application Serial No. 511,870 and Patent No. 3,229,775, the tubular rotatable valve passes through a central circular passage or bore through the hammer and the hammer reciprocates up and down on the valve. Pressure sticking and frictional resistance to rotation of the valve in the hammer cause nonuniformity in the timing relationship between the hammer and valve and a great loss in the power output of the percussive unit. This invention relates to an improvement in the rotatable valve which overcomes a cause of pressure sticking and frictional resistance and to a percussive unit using this improved rotatable valve so as to provide an improved percussive unit.

Other advantages and objects of this invention will become apparent lby reference to the accompanying drawings, appended claims and following specification.

In the drawings:

FIGURES 1 and 2 are elevational, fragmented, partial cross-sectional views showing the internal construction of -one embodiment of the percussive mechanism described herein.

FIGURE 3 is an elevational, cross-sectional view of the hammer of FIGURE 1.

FIGURES 4 and 5 are top and bottom views, respectively of the hammer of FIGURE 3.

FIGURE 6 is a side, elevational, cross-sectional view of the hammer of FIGURE 3.

FIGURE 7 is a fragmented, elevational, cross-sectional view of the rotatable valve of FIGURE 1.

FIGURE 8 is a fragmented, side, elevational view of the rotatable valve of FIGURE 7.

FIGURE 9 is a top, cross-sectional view taken at 9-9 in FIGURE 8 showing the symmetry of the exhaust and inlet ports in the valve.

FIGURE 10 is a top, cross-sectional view taken lat 10-10 in FIGURE 1 and showing one embodiment of the Huid-actuated rotary drive means for the percussive mechanism.

FIGURE 11 is an isometric view of a single-acting cam as shown in FIGURE 1.

FIGURES 12, 13, and 14 are a front, side and top view, respectively, of a follower means as shown in FIGURE 1.

FIGURE 15 is the mathematical ellipse best tting the ported section of the valve illustrated in FIGURE 9.

FIGURE 16 is an enlarged diagram depicting the elliptical section of the valve in a circular hammer bore and illustrating some of the features of the elliptical section.

Briefly, this invention provides a ported tubular, rotatable valve for a fluid-operated percussive unit. The valve has an elliptical or oval shape in the vicinity of the ports. More sepcifically, the tubular rotatable valve has two high pressure power fluid inlet ports located on opposite sides of the valve and spaced approximately degrees on center so that the high pressure power Huid inlet ports have essentially the same axis of symmetry. The tubular rotatable valve also has two power uid exhaust ports. The exhaust ports are located on opposite sides of the valve and spaced approximately 180 degrees on center. The exhaust ports are located on opposite sides of the from the symmetrically spaced high pressure power fluid inlet ports. The section of the valve in the vicinity of these ports has an exterior surface which is between the ports. This exterior surface has a conguration which 3 gives the valve a generally elliptical cross section with the minor axis of the elliptical surface centering on the high pressure power fluid inlet ports and the major axis of the elliptical surface centering on the power fluid exhaust ports. Preferably, the valve passes through a cent-ral circular passage or bore through a hammer. The inside diameter of the hammer bore is greater than the maximum width fo the elliptical section of the valve to form a diametral clearance. The difference between the major and minor axis of elliptical surface is at least equal to this diametral clearance, and more preferably, this difference is at least twice the diametral clearance. In another embodiment, the diametral clearance is at least 0.002 inch and the difference between the major and minor axis of the elliptical surface is at least 0.002 inch, and more preferably, the difference is at least 0.004 inch.

In another embodiment of this invention, a portion of the exterior surface of the valve between the inlet and exhaust ports are in the shape of two eccentric arcs that have offset centers of curvature. The two arcs are sym- :metrically spaced on opposite sides of the valve and have the same axis of symmetry as the axis of symmetry of the high pressure power fluid inlet ports. Preferably, this valve passes through a central circular passage or bore through a hammer. The inside diameter of the hammer bore is greater than the maximum width of the ported section of the valve to form a diametral clearance and the distance between the center of curvature of the two eccentric arcs is at least twice this diametral clearance, and more preferably, this distance is at least four times this diametral clearance. In another embodiment, the diametral clearance is at least 0.002 inch and the distance between the centers of curvature of the two eccentric arcs is at least 0.004 inch and, more preferably, this distance is at least 0.008 inch.

In other embodiments of this invention, the special tubular valve is combined with a rotary drive means adapted to rotate the valve relative to the hammer. The valve is also combined with a motion conversion means adapted to employ the reciprocating movement of the hammer to coordinate the axial position of the hammer with the circumferential rotative position of said ports in the valve during at least a portion of the reciprocating cycle of the hammer. More specifically, the motion conversion means may be a cam and follower means, the cam having at least one curved surface positioned to contact the follower during at least a portion of the reciprocating cycle of the hammer. In still a yfurther embodiment, there is a flow restrictive means in the valve Ibetween the high pressure power fluid inlet ports and the power fluid exhaust ports. The flow restrictive means is adapted to continuously pass a small amount of power fluid lfrom an upper inlet section to a Ilower outlet section of the valve.

In a more specific embodiment, the special tubular valve and the above-mentioned other elements are cornbined with the elements of a specific percussive unit and with at least one elongated rod-like spring member mounted above the hammer of the percussive unit. The rodlike spring member is at least twelve times as long as it is wide with its longitudinal axis substantially parallel to the longitudinal axis of the percussive unit. The lower end of the rod-like spring member is positioned to co-ntact the upper end of the hammer and receive energy from the hammer during at least part of the upward movement of the hammer and to return a portion of this energy during a first part of the downward movement of the hammer.

The word rotatable as used herein in relation to the tubular valve includes a tubular valve that rotates back and forth in an arc of one revolution or less as well as one that rotates in only one direction.

More specifically, in FIGURES l and 2, there is shown one embodiment of a percussive unit Ibuilt in accordance with the present invention. The percussive unit includes tu'bular housing 11 whose upper end is removably connected to a rotary drill string (not shown) which con'- ducts power fluid to the percussive unit.

Slidably telescoping into the lower end of housing 11 is anvil 13 having an upper anvil surface and which is capable of limited longitudinal movement within the housing. Longitudinally traversing the anvil is central anvil bore passage 15 whose lower end (not shown) communicates in the usual fashion with fluid discharge passages in a drill bit. In the usual manner, the anvil is designed to rotate with housing 11.

In housing 11 above anvil 13 is piston-like hammer 17 which is slidably mounted within housing 11 to undergo reciprocal, longitudinal, or up and down movement and periodically impact anvil 13. The hammer has upper surface 19 and lower surface 21. In the illustrated embodiment the lower portion of hammer 17 is of smaller diameter than the upper portion, thereby lforming shoulder 23 at a point intermediate of the ends of the hammer.

As shown in FIGURES 1 and 3 through 6, longitudinally traversing hammer 17 is central bore 25. Communicating with this central bore passage is at least one upper or first hammer passage 27 which also communicafes with the upper surface or end of the hammer and the interior of housing 11 above the hammer. As shown, first hammer passages 27 are two vertical grooves leading to the top of the hammer which are spaced 180 degrees on center and substantially symmetrically around the circumferenceof central bore 25. These passages extend deep enough into the bofe of the hammer to communicate with ports in a rotatable valve and, during op= eration lof the percussive unit, these first hanr'rieijpsd sages conduct power fluid to and from the cylinder charnber above the hammer as hereafter set forth.

Also communicating with the central bore passage in the hammer is at least one lower or second hammer passage 29 which in turn communicates with the lower surface or end of the hammer by way of shoulder 23. As shown, second hammer passages 29 include a groove in the wall of the central bore of the hammer, two bores which pass through the walls of the hammer and two vertical grooves in the outer surface of the hammer. These passages are spaced 180 degrees on center and substantially symmetrically around the circumference of cend tral bore 25. These passages are also spaced degrees on center from first hammer passages 27. Second hammer passages 29 are located in the hammer bore to communicate with ports in a rotatable valve, and, during operation of the percussive unit, these second hammer passages conduct power uid to and from the lower end of the hammer as hereafter set forth.

Extending longitudinally through the central bore in hammer 17 and into the central bore of the anvil is a rotatable tubular valve 31 which, as illustrated, has a central passage which is open at its lower end and which near its upper end has inlet openings 33. The tubular valve is free to rotate substantially independently of housing 11 and the anvil and bit.

Near the top of the housing and tubular valve is rotary drive means 35 which is connected to or made a part of .tubular valve 31. The rotary drive means will be hereafter described in more detail.

Below rotary drive means 35 and inlet openings 33 in tubular valve 31 are two high pressure power fluid inlet ports 37 through the wall of the tubular valve which, when the drill -bit is on bottom and the valve is rotated, alternately communicate with the first and second hammer passages to conduct power fluid to above and below the hammer. The two inlet ports are spaced substantially symmetrically Iaround `and through the circumference of the tubular valve as shown in FIGURES l and 7 through 9, wherein the two inlet ports are degrees on center and have the same axis of symmetry. The spacing and location of these ports will be discussed in further detail` Through the wall of the tubular valve are also two power fluid exhaust ports 39, which, when the drill bit :is on bottom and the valve is rotated, alternately communicate with the first and second hammer passages to conduct power uid from :above and below the hammer. The two exhaust ports are spaced substantially symmetrically around and through .the circumference of the tubular valve. As shown, the two such exhaust ports are 180 degrees on center and are spaced 90 degrees on center from each inlet port.

Normally high pressure power fluid inlet ports 37 and power fluid exhaust ports 39 will overlap -a common horizontal or lateral plane; however, lthis location is not essential to this invention. In any case, the valve in the vicinity of these ports has an exterior surface which is between the ports and it is this surface which is important to this invention as hereafter explained. It should also be realized that the ported section of the tubular valve will have maximum width less than the internal diameter of central bore in hammer 17. In other words, there will be a clearance between the exterior surfaces of the ported secti-on of the valve and the wall of the hammer bore. This clearance is too small to be noticed in the drawings; however, this clearance is illustrated in FIGURE 16 and the pertinence of this clearance and the surfaces between the ports will be discussed hereafter in greater detail.

In tubular valve 31 separating high pressure power fluid inlet ports 37 from power fluid exhaust ports 39 is flow restrictive means 41 which effectively divides the central passage through tubular valve 31 into two sections with the section above the flow restrictive means being la power fluid inlet section and the section below the flow restrictor being a power uid exhaust section. For illustrative purposes only, the wall of the valve and flow restrictor is equipped with grooves that give greater vertical length to inlet ports 37 and exhaust ports 39.

Flow restrictive means 41 has bleed port 43 which provides greater control of the conductivity of the percussive unit and permits power fluid to escape from the power inlet section of the tubular valve to the exhaust section thereby assuring a continuous stream of fluid for cleaning and flushing the drill bit.

As stated previously, through the wall of tubular valve 31 rabove power Huid inlet ports 37 and below rotary drive means 35 are inlet openings 33 which communicate with power fluid passageway 45 in casing 11 which conducts power fluid from rotary drive means 35 to tubular valve 31.

Circumscribing and supporting tubular valve 31 is bearing 47 which permits easy rotation of the valve. Preferably bearing 47 will be a ball, thrust bearing.

Return now to rotary drive means 35. This rotary drive may be any form of fluid-actuated rotary device or fluid reaction motor connected to and suitable for rotating tubular valve 31 relative to the hammer and may or may not be `an integral part thereof. As shown in FIGURES 1 and 11, rotary drive means 35 is of the jet type and includes hollow central chamber 49 which communicates with incoming power uid from a drill string attached to the upper end of housing 11. Extending outward from central chamber 49 are jet passages 51 and 53 which terminate in jet openings 55 and 57.

The top of central chamber 49 forms a rotatable seal with insert 59 in housing 11. Power fluid passing into central chamber 49 passes to jet openings 55 and 57 located on opposite sides of the rotary drive means so that fluid emitting from these jet openings will cause the rotary drive means to rotate.

Above hammer 17, housing 11 provides a chamber for an elastic rebound cushion which, during a last portion of the upward travel of the hammer, receives and stores energy from the hammer causing the hammer to decelerate and which, during a first part of the downstroke of the hammer, returns a portion of the stored energy to the hammer.

Preferably, the elastic rebound cushion is at least one compression rod spring. As shown in FIGURE l, there are two elongated compression bar or rod-like springs 61 and 63 mounted in housing 11 above hammer 17. These rod-like springs are positioned and adapted to be loaded axially when struck by upper surface 19 of the hammer; consequently, rod-like springs 61 and 63 have lower ends 65 and 67, respectively, which are positioned above the hammer to be squarely contacted and receive energy from the hammer. The sides of each rod-like spring are supported laterally by housing 11 or other support to prevent buckling; consequently, the rod-like springs act only in axial compression. The length of each rod-like spring is at least twelve times as great as the average maximum width of the main body of the rod, and much greater length to width ratios are desired. The spring coefficient for the rod-like spring will normally be greater than 5,000 pounds per inch and spring coefficients lmuch greater than this are preferred.

The ported, rotatable tubular valve is combined with motion conversion means adapted to employ the reciprocating or longitiudinal movement of the hammer to coordinate the axial motion and position of the hammer with the circumferential or rotative position of the ports in the valve during at least a portion of the reciprocating cycle of the hammer. In other words, the motion conversion means converts longitudinal motion into rotary motion. rl`his motion conversion means may be a cam and follower arrangement with the cam having at least one curved surface adapted to contact the follower during at least a portion of the reciprocating cycle of the hammer. In a cam and follower arrangement, either the cam or the follower may be in Contact with or connected to the tubular valve. As illustrated in FIGURES 1 and 11, single-acting cam 69 is embedded in the upper end of hammer 17 surrounding tubular valve 31 so that the cam moves up and down with the hammer. The top v of single-acting cam 69 has single-acting cam surface 71 which has the same general form as the relative motion between the valve and hammer. Since the tubular valve has two power fluid inlet ports and two power tiuid eX- haust ports spaced symmetrically around the valve, there is a hammer cycle every degrees of rotation of the valve or two cycles per valve revolution; consequently, there are two single-acting cam surfaces of the same shape with each surface covering 180 degrees of the cam cylinder.

Above `single-acting cam 69 is a follower means which is a member, such as a pin, roller or second cam surface, adapted to cooperate with longitudinal movement of single-acting cam 69 during at least a portion` of the up and down strokes of the hammer by riding on singleacting cam surface 71 whenever the ports in the tubular valve are not in proper position relative to the -up or down position of the hammer. Thus, single-acting cam surface 71 will either .speed up or slow down tubular valve 31 whenever ports 37 and 39 are out of position relative to the position of the hammer. As shown, the follower means is pin or roller 73 which traverses tubu lar valve 31 extending outward through follower slot 75. The follower roller extends beyond the wall of the tubular valve by a distance sufficient to allow the follower roller to be vertically aligned with single-acting cam surface 71.

Preferably, either the follower or the cam should be shock mounted to allow either one to slip or give way if there is a malfunction which causes a mismatch between the up and down motion of the hammer and the rotational movement of the tubular valve. This shock mounting helps to prevent failure or breakage of the percussive unit. One method of shock mounting the follower roller 73 is shown in FIGURES l and 12 through 14. The follower roller is affixed to cam follower bushing 77 which in turn is connected to cam follower bushing return spring 79?. It will be also noted that follower slot 75 is elongated longitudinally along the length of tubular valve 31. In this manner, whenever there is a mismatch between cam and follower, follower roller 73 will slip upward in follower slot 75 causing cam follower bushing 77 to compress cam follower bushing return spring 79 which will return the follower to its original position when the mismatch is corrected or the hammer moves downward out of contact with the follower.

The following description of the operation of the percussive unit of FIGURE l will aid in understanding the elliptical portion of the valve. First, note that tubular valve 31 is rotated thereby rotating high pressure power fluid inlet ports 37 and power fluid exhaust ports 39. As shown, there are two inlet ports and two exhaust ports with the two inlet ports being 180 degrees on center and having the same axis of symmetry. The two exhaust ports are also 180 degrees on center with each inlet port spaced 9() degrees on center from each exhaust port. As shown, all of the ports are of substantially the same length and size and, when rotated, are adapted to communicate with the same cylindrical area; however, the relative sizes of the ports may be varied to fit the operating characteristics of the system. Using the configuration shown, a complete yhammer cycle occurs every 180 degrees of rotation of the valve.

As to the upper and lower passages in the hammer, note that upper or first hammer passages 27 and lower or second hammer passages 29 reciprocate up and down with hammer 17. As far as the operation of the percussive unit is concerned, these passages do not rotate when opening and closing to the ports in the rotating valve. A hammer cycle includes one upstroke and one downstr-oke per every 180 degrees of rotation of the tubular valve.

It should also be noted that the basic components of the percussive unit are a ported rotating valve for supplying power fluid to and removing power fluid from the ends of the hammer, a reciprocating hammer with upper and lower passages, a cam and follower, and a rebound spring above the hammer.

At the start of a hammer cycle, the hammer rests on anvil 13, inlet ports 37 are just opening to lower the hammer passages 29 and exhaust ports 39 are just opening to upper hammer passages 27. Power fluid enters rotary drive means 35 via insert 59 and passes outward through jet passages 51 and 53. The power fluid emits from jet openings 55 and 57 causing the rotary drive means to rotate tubular valve 31. Power fluid passes downward in passageway 45 through inlet openings 33 and into inle-t section of tubular valve 31. Power fluid inlet ports 37 communicate with lower or second hammer passages 29 which in turn communicate with the underside surfaces of the hammer. The power fluid pressure acting on the underside surfaces of the hammer causes the harnmer to accelerate upward in its up or return stroke. At the same time, powe-r fluid above the hammer is exhausted through upper or first hammer passages 27, power fluid exhaust ports 39 and the exhaust section of tubular valve 31 where the exhausted fluid is conducted through anvil passage to the drill bit.

Hammer 17 accelerates upward causing single-acting cam surface 71 to rise to a point where the cam surface could contact cam follower 73. Whether or not cam follower 73 contacts cam surface 71 will depend on the angular or rotative position of tubular valve 31. If the valve is rotating either too fast or too slow relative to the position of the hammer, single-acting surface 71 will contact the cam lfollower causing the follower to follow the cam surface thereby coordinating the axial motion and position of the hammer with the angular or circumferential rotative position of inlet ports 37 and exhaust ports 39 in the tubular valve.

Hamme-r 17 continues to accelerate upward in its upstroke as long as power fluid passes from inlet ports 37 to second hammer passages 29 and the underside of the hammer. At the desired moment of upward hammer travel and valve rotation, power fluid inlet ports 37 ro tate out of communication with second hammer passages 29 cutting off the flow of power fluid to the underside of the hammer. The exact point in the upward trave-l of the hammer when the valve shuts off the flow of power fluid will be adjusted to the compressibility of the power fluid, to the type of rebound cushion above the hammer, and to other factors.

In the embodiment shown, at the same time as power fluid inlet ports 37 rotate out of communication with second hammer passages 29, power fluid exhaust ports 39 rotate out of communication with first hammer passages 27; however, it should be realized that the size or shape of these ports could be Varied to adjust the timing of these occurrences to suit the powe-r fluid and other operating conditions. For example, if the power fluid were an incompressible liquid, it would be desirable to permit the fluid above the hammer to escape through power fluid exhaust ports 39 until approximately the moment that the hammer reaches its maximum upward travel as he-reafter described.

When power fluid inlet ports 37 rotate out of communication with the second hammer passages, the hammer continues to travel upward decelerating and compressing springs 61 and 63 until the upward energy of the hammer is absorbed by these springs and the drill string. Preferably, the hammer will contact the springs at the same moment as the flow of power to the underside of the hammer is cutoff.

Before the hammer reaches its upward peak, tubular valve 31 rotates far enough for power fluid inlet ports 37 to communicate with upper or first hammer passages 27. High pressure power fluid passes upward via passages 27 to exert its pressure on the upside surfaces of the hammer. This pressure causes some deceleration of the hammer and, once the hammer reaches its peak upward travel, causes the hammer to accelerate downward.

In the embodiment shown, at the same instant as power fluid inlet ports 37 rotate into communication with first hammer passages 27, power fluid exhaust ports 39 rotate into communication with lower or second hammer passages 29. This allows fluid beneath the hammer to eX- haust through second hammer passages 29 and power fluid exhaust ports 39 and into anvil bore 15.

The hammer accelerates downward leaving springs 67 and1f69 and impacting the anvil. The cycle then repeats itse It should be noted that during a major and critical portion of each hammer cycle when the hammer is changing from its upstroke to its downstroke, single-acting cam surface 71 acts as a limit or control on the position of cam follower 73 to assure that inlet ports 37 and eX- haust ports 39 open and close to first and second hammer passages 27 and 29 at the optimum moment as determined by the position of the hammer.

From the foregoing discussion of the operation and cycle of the hammer and rotatable tubular valve, it should be apparent that the timing and smoothness of operation of the valve affects the entire operation of the percussive unit, especially the impact energy and frequency, power fluid consumption and starting of the valve. One problem encountered with a rotating valve is irregular speed of rotation. One cause of irregular valve movement is friction. Some of the friction can be overcome by providing adequate clearance between the tubular valve and hammer bore and keeping the valve and hammer bore in close alignment; however, it is diicult to precisely center the valve in the hammer bore and to maintain this alignment as the parts of the percussive unit wear. In addition, there occurs the problem of friction due to uneven distribution of power fluid in the clearance between the valve and hammer bore. In order to overcome the friction caused by uneven pressure distribution in the hammer-valve clearance, it was proposed in copending application Serial No. 511,870 that the number of ports be increased and that the ports be symmetrically arranged. This circumferential or radial symmetry better balances the radial forces caused by the pressure of the power fluid. In this invention, this symmetry is accomplished by providing two power fluid inlet ports 180 degrees on center and two power fluid exhaust ports 180 degrees on center and spaced so that each exhaust port is 90 degrees on center from an inlet port. This approach to eleviating pressure sticking of the valve is modestly successful; however, symmetry alone does not prevent pressure sticking of the valve. It was found that the valve tends to stick whenever the valve is not centered in the hammer bore. It was also discovered that pressure sticking is caused by an eccentric side force on the valve. The causeI of the eccentric side force will be better understood by reference to the following description; however, it may be pointed out in advance that it was discovered that this eccentric side force is due to a change in convergence and divergence of the flow paths leading from the high pressure inlet ports to the exhaust ports in the clearance between the valve and hammer bore when the valve is not centered in the hammer bore.

In brief, pressure sticking of a tubular rotating valve is overcome herein by slightly altering the configuration of the outer surface of the valve so as to change the power tluid pressure distribution in the clearance so that the power fluid pressure will create a net force toward the center of the hammer bore whenever the valve becomes off center. This is accomplished by providing the valve with a slightly elliptical outer surface in the vicinity of and between the inlet and exhaust ports with the minor or short axis of the elliptical portion of the valve centering on the high pressure inlet ports.

This invention, therefore, refers to an exterior surface of a rotatable valve in the vicinity of or lbetween two high pressure inlet ports and two exhaust ports. A portion of this surface is sufficiently elliptical or oval in shape that when the valve is held against the wall of a circular passage having a diameter greater than the maximum width of the elliptical portion -of the valve, the flow passages between the side of the valve and the nearest wall -of the circular passage are more convergent in the direction of fluid flow than the flow passages on the opp-osite side of the valve furthest from the w-all of the circular pass-age.-

It is not necessary that the elliptical or oval portion of the valve have a cross section in the shape of a mathematical ellipse. Whether or not the cross section of the valve follows a mathematical ellipse or is only approximately elliptical, the key factor is the ratio of the convergence of the flow paths on the side of the valve nearest the hammer bore to the convergence of the flow paths on the opposite side of the valve.

Normally, the broadly elliptical or oval shape of the valve is generated by or deviates from a circle through either elongation or extension of one axis to form a major or longer axis, or compression of the other axis to form a minor or shorter axis, or both. In any event, the curvature of a substantial portion of the surface between the high pressure power uid inlet ports and the power fluid exhaust ports should -be sufficiently elliptical or oval in pattern that a mathematical ellipse best fitting the pattern of the surface can be established so that the difference in length between the major and minor axes of this mathematical ellipse may be reason-ably established. rI`his difference is important in that it may be related to the diametral clearance of the ported section of the valve and used to determine if the ported section of the valve is sufficiently elliptical for the power fluid pressure distribution to prevent pressurel sticking of the valve. For example, FIGURE shows the mathematical ellipse best fitting the valve cross section shown in FIGURE 9. This ellipse has minor axis S and major axis L. The

10 relationship between the elliptical section of the valve and the diametral clearance is stated and illustrated in another manner in relation to FIGURE 16 which will be hereinafter described in more detail.

The elliptical section of the valve distributes the power fluid pressure in the clearance between the valve and the walls of the hammer lbore in a way that the valve will be forced toward the center of the hammer -bore whenever the valve becomes off centered. How and why the elliptical section of the valve accomplishes this centering is -best illustrated by-referring to FIGURE 16.

In FIGURE 16, there is illustrated circular portion 81 of a circular valve and elliptical portion 83. Symmetrically spaced around elliptical portion 83 are two high pressure inlet ports 37 and two exhaust ports 39. The inlet ports are centered on the minor or short axis of the elliptical portion and the exhaust ports are centered on the major or long .axis of the elliptical portion. The circular portion and elliptical portion have a maximum physical width less than the diameter of the hammer bore. The difference between the diameter ofthe hammer bore and the maximum width of the elliptical section of the valve is herein referred to as the diametral clearance, which in FIGURE 16 is B.

In FIGURE 16, the valve 4occupies an off center position in hammer bore 25 with one of the high pressure inlet ports closest to the wall of the hammer bore.

By referring only to circular portion 81 of FIGURE 16, one can see that if the inlet ports were in a circular valve, the flow paths in the clearance from the near side inlet port to the exhaust ports would be divergent. On the other side of the circular portion, the flow paths from the opposite inlet port to the exhaust ports would be convergent. There are, therefore, two convergent paths and two divergent paths. All four ow paths would be essentially equal in length due to the symmetry of the four ports. The paths would also either `converge or diverge uniformly from an inlet port to an exhaust port and exit abruptly into a relatively large exhaust opening.

Thus, there would be a net force exerted by the power fluid in each passage. This force can be treated as a force normal to the boundary of each ow path. The magnitude force is dependent on the convergence or divergent of the flow path. The net force increases with increasing convergence of the flow passage in the direction of flow. When a circular valve occupies the olfcentered position just described, the most convergent paths are on the side of the valve having the most clearance while the flow paths on the side of the valve nearest the hammer bore are divergent. As a result, there would exist a large net force pushing the valve toward the side where the valve is closest to the wall of the hammer bore. This causes pressure sticking of a circular valve.

In this invention, t-he surface 4of the valve between or in the Vicinity of the ports is made elliptical with the high pressure inlet .ports centering on the minor axis. An example of an elliptical surface is illustrated in FIGURE 16 by elliptical portion 83. By -referring to this elliptical portion, one can see that the flow pat-hs from the near side inlet port to the exhaust ports are convergent. The corresponding flow paths on the other side of .the circular portion are also convergent. Thus, the elliptical shape has made all flow paths convergent. This alone improves the operation of t-he valve; however, it is preferred that the ow paths on the side of the valve nearest the wall of the hammer be more convergent than the opposite How pa-ths. The degree of convergence is the ratio of the inlet area of a flow path over the -outlet area of the ow path. In the elliptical shape of elliptical portion 83, the flow paths on the side of the valve nearest the wall of the hammer bore are more convergent than the llow paths on the opposite side of the valve. It has previously been noted that the force increases with increasing convergence; consequently, there will be a net force to push the off-centered elliptical valve back toward the cen-ter of the hammer bore. In other words, the surface is sufficiently elliptical 'that whenever the valve occupies an off-center position, the iiow paths in the clearance from an inlet port to t-he l'exhaust ports on the side of the valve closest to the wall of the hammer bore are convergent `and are more convergent than the fiow pa-ths on the opposite side of the valve. As stated previously, increasing the convergence of the paths increases the net force; consequently, the force is greatest on the sid-e of the valve nearest the Wall of the hammer bore. This net force pushes the valve back toward t-he center of the hammer bore, thereby preventing pressure sticking of the valve. The magnitude of the centering force will change with the nature of the flow, e.g., compressibility, viscosity and turbulence; however, the quality or direction of the radial force is determined primarily by the geometry of the flow passages.

It has been shown that in the case of the off-center circular valve, there are two flow paths that were convergent and two that were divergent. If the shape of the exterior surface of a circular valve is gradually changed toward an elliptical shape with the high pressure inlet ports centering on the minor axis, t-he divergent paths become less divergent and eventually there is a crossover point where all of the flow paths are convergent. Even though the crossover point is reached, it is desirable to make the surface more elliptical so tha-t the flow paths in the clearance between the side of the valve nearest the wall of the hammer bore will be more convergent than the flow paths on the opposite side of the valve so that the net eccentric force will always be toward the center of the hammer bore. In order to assure that the flow paths on the side of the valve nearest the wall of the hammer bore are sufficiently more convergent than the opposite flow paths to create a net force toward the center of the hammer bore, minimum limits can be placed on .the elliptical shape of the valve. These minimum limits -will be lbased on rst, a practical basis, and second, a more mathematical basis. On either "basis, the minimum limits are dependent on the amount of diametral clearance and the amount of elongation, or compression, or both, of the elliptical shape.

As mentioned above, the minimum limits on the elliptical shape are dependent on -t-he diametral clearance between the po-rted sec-tion of the valve and the hammer bore. It should be emphasized that the clearance under consideration is diametral clearance and not radial clearance. In other words, the clearance is the difference between the inside diameter of the hammer bore and the maximum width of the ported elliptical section of the valve. It should be kept in mind that, as a practical matter, the clearance on sides of the elliptical section of the valve must be kept small enough to prevent excessive leakage of power fluid between the inlet and exhaust ports; however, it has been found that in percussive units of this -type some clear-ance is required. This is in contrast to the very small clearances found in many hydraulic sys-tems wherein clearances of the order of 0.0001 inch are common. It has been found that a diametral clearance of at least 0.002 inch is desirable; therefore, on a practical basis the difference between the major and minor axes of the elliptical portion of the'valve should be at least 0.002 inch, and preferably at least 0.004 inch.

On a mathematical basis, the difference between t-he major and minor axes of the ellipse must be at least equal to the diametral clearance of the elliptical-portion of the valve in the hammer bore, and it is preferred that the difference between the major and minor axes be at least twice the diametral clearance. Larger values will improve the centering force or effect on the valve at the expense of greater leakage of power fluid through the clearances so tha-t a compromised value must be accepted. For example, agood compromised value is ve times the diametral clearance.

As stated previously, it is not necessary that the oval or elliptical cross section of the valve be a mathematical ellipse. Normally, it will be more convenient to generate the approximately elliptical section by forming eccentric arcs on the valve as illustrated in FIGURE 16. The eccentric arcs have cen-ters of curvature offset from each other and from the center of the valve. The distance between the centers of curvature of the two arcs is related to the diametral. clearance to accomplish the centering effects previously discussed.

In FIGURE 16, as -a practical machine operation, the new valve elliptical cross 4section is formed by grinding or milling two eccentric arcs on the sides of lthe valve having the inlet ports so that the minor or short axis of the valve will center on these inlet ports. The two eccentric arcs are symmetrically spaced on opposite sides of the valve and have an axis of symmetry that is the same as the axis of symmetry of the high pressure power uid inlet ports. The depth of cut of each arc and radius of curvature R of the `arcs are controlled by the desired amount of compression of the sides of the valve and by points D on the valve where the cuts are to run out of the original surface of the valve. Each runout point depends on the width of the inlet and exhaust port-s. In the example illustrated, these points are 67.5 degrees from the center of the high pressure port. Radius R of each arc will be determined by distance from the center of curvature of the arc to the point of runout. The center of curvature of the arc is located on an axis of symmetry through the centers of the high pressure inlet ports and center of the valve with the center of curvature of the arc being offset from the center of the valve by offset distance A/ 2 established by determining the amount of offset required to make the surface sufficiently elliptical to kcause fiow paths in the clearance to be convergent enough to center the valve. I-t has been found that the center of curvature of each arc must be offset from the center of the valve by a distance Iat least equal to diametral clearance B between the ported section of the valve and the hammer bore, and an offset of at least twice the diametral clearance is preferred. Since the centers of curvature of the two arcs are equally spaced from the center of the valve, total offset distance A between the two centers of curvature of the arcs will be at least twice diametral clearance B with at least four times the diametral clearance being preferred. The centers of curvature of the arcs are set to meet these limits and the necessary cuts taken.

A rotating valve having the above-described features performed well when tested with both liquid and gaseous power fiuid.

In the foregoing discussion of the elliptical section of the valve, for the sake of simplicity, the effects of and on the hammer passages have been ignored.

While the invention has been described in connection with certain specific embodiments thereof, it will now be understood that further modifications will suggest themselves to those skilled in this art, and it is intended to cover such modifications as fall within the scope of the foregoing description and appended claims. For example, it should be recognized that the first and second hammer passages or the power fluid inlet and outlet valve ports may be varied in number, location and size to adapt the tool to different modifications. It should also be recognized when the words ports and openings have been used herein in conjunction with the high pressure power fiuid inlets to the hammer passages and with the power fluid exhaust outlets, the word port or opening 1s not limited to single openings. Each port or opening may be one or more fiow passages, slots or holes provided that they perform together to essentially act at the same point in the cycle of the hammer cycle to pass power fluid to the ends and passages of the hammer.

I claim:

1. In a power fluid-operated percussive unit of a rotary percussive drilling system having a hammer and a drilling means yand wherein the percussive unit and the drilling means are rotated and the drilling means is impacted against the earth `by axial cyclic reciprocation of the hammer, and the power fluid is passed through a tubular, rotatable valve having two high pressure power fluid inlet ports located in said valve in a position to alternately pass power fluid to passages communicating with the opposite ends of said hammer as said valve is rotated, and two power fluid exhaust por-ts located in said valve'in a position to alternately pass power fluid from said passages communicating with said opposite ends of said hammer to a power fluid exhaust passage as said valve is rotated, the improvement comprising said two high pressure power fluid inlet ports being located on opposite sides of said valve and approximately 180 degrees on center so that said high pressure power fluid inlet ports are spaced substantially symmetrically around the surface of said valve, said two power fluid exhaust ports being located on opposite sides of said valve, said two power fluid exhaust ports being approximately 180 degrees on center from each other and approximately 90 degrees on enter from said high pressure power uid inlet ports and the exterior surface of said valve `between said high pressure power fluid inlet ports and said power fluid exhaust ports having a generally elliptical configuration with said high pressure power fluid inlet ports centering on the minor axis of said generally elliptical surface and said power fluid exhaust ports centering on the major axis of said generally elliptical surface.

2. The improvement of claim 1 wherein the rotatable tubular valve is connected to rotary drive means adapted to rotate said valve relative to the hammer.

3. The improvement of claim 1 wherein there is a flow restrict-ive means in said Valve bet'ween the high pressure power fluid inlet port-s and the power fiuid exhaust ports, said flow restrictive means adapted to continuously pass a small amount of power fluid from an upper inlet section to a lower outlet section of said valve.

4. The improvement of claim 1 wherein there is motion conversion means adapted to employ the reciprocating movement of said hammer to coordinate the axial position of said hamm-er with the circumferential rotative position of sa-id ports in said valve during at least a por tion of the reciprocating cycle o-f said hammer.

S. The improvement of claim 4 wherein the motion conversion means is a cam and follower means, said cam having at least one curved surface positioned to contact said follower during at least a portion of the reciprocating cycle of said hammer.

6. The improvement of claim 1 wherein the valve passes through a central circular bore through the hammer, said central circular bore having an inside diameter greater than the maximum width of the elliptical section of said valve to form a diametral clearance, and the difference between the maj-or and minor axes of said elliptical surface being at least equal to said diametral clearance.

7. The improvement of claim 6 wherein the rotatable tubular valve is connected to rotary drive means adapted to rotate said valve relative to the hammer.

8, The improvement of claim 6 wherein there is a flow restrictive means in said valve between the high pressure power fluid inlet ports and power fluid exhaust ports, said flow restrictive means adapted to continuously pass a small amount of power fluid from an upper inlet section to a lower outlet section of said valve.

9. The movement of claim 6 wherein there is motion conversion means adapted to employ the reciprocating f movement of said hammer to coordinate the axial position of said hammer with the circumferential ro-tative position of said ports in said valve during at least a portion of the reciprocating cycle of said hammer.

10. The improvement of claim 9 wherein the motion conversion means is a cam and follower means, said cam having at least one curved surface positioned to contact said follower during at least a portion of the reciprocating cycle of said hammer.

11. The improvement of cla-im 6 wherein the difference between major and minor axes of said elliptical surface is at least twice the diametral clearance.

12. The' improvement of claim 11 wherein the rotatable tubular valve is connected to rotary drive means adapted to rotate said valve relative to the hammer.

13. The improvement of claim 111 wherein there is a flow restrictive means in saidvalve between the high pressure power fluid inlet ports and power fluid exhaust ports, said flow restrictive means adapted to continuously pass a small amount of power fluid from an upper inlet section to a lower outlet section of said valve.

14. The'impnovement of claim 11 wherein there is motion conversion means adapted to employ the reciprocating movement of said hammer to coordinate the axial position of said hammer with the circumferential rotative position of said ports in said valve during at least a portion of the reciprocating cycle of said hammer.

15. The improvement of claim 1'4 wherein the motion conversion means is a cam and follower means, said cam having at least one curved surface positioned to contact said follower during at least a portion of the reciprocating cycle of said hammer.

16. The improvement of claim 1 wherein the valve passes through a central circular bore through the hammer, said central circular bore having an inside diameter greater than the maximum width of the elliptical section of said valve to form a diametral clearance of at least 0.002 inch and the difference between the major and minor axes of said elliptical surface being at least 0.002 inch.

17. The improvement of claim 16 wherein the difference between the major and minor axes of said elliptical surface is at least 0.004 inch.

18. In rotary percussive earth drilling wherein a percussive unit and an eartlh drill are rotated and the drill is impacted against the earth by axial cyclic reciprocation of a hammer within the percussive unit, a power fluid-operated peroussive unit comprsing an elongated casing, an anvil surface sl-idaibly mounted in the lower end of said casing, a piston-like hammer slidably mounted for reciprocating up and down movement in said casing above said anvil surface, said hammer having an upper surface and first hammer passages communicating with said upper surface, said hammer having a lower surface and second hammer passages communicating with said lower surface, a rotatable tubular valve passing longitudinally through a central bore throrugh said hammer, said tubular valve having a valve inlet passage adapted to pass fluid into said valve and a valve outlet passage adapted to pass power fiuid from said valve, two high pressure power fluid inlet ports located in said valve in a position to alternately pass said power fluid from said valve inlet passage to said first hammer passages and to said second hammer passages as said valve is rotated, two power fluid exhaust ports located in said valve in a position to alternately pass said power fluid from said first and said second hammer passages to said valve outlet passage as said valve is rotated, said two high pressure power fluid inlet ports being located on opposite sides of said valve and approximately 180 degrees on center so that said high pressure power fluid inlet ports are spaced substantially symmetrically around the surface of said valve, said two power fluid exhaust ports being located on opposite sides of said valve, said two power uid exhaust ports being approximately 180 degrees on center from each other and approximately degrees on center from said high pressure power fluid inlet ports, and the exterior surface of said valve between said high pressure power fluid inlet ports and said power fluid exhaust ports having a generally elliptical configuration with said high pressure power fluid inlet port-s centering on the minor axis of said generally elliptical sur-face and said power fluid exhaust ports centering on the major axis of said generally elliptical surface.

1-9. The percussive unit of claim 18 wherein there is at least one elongated rod-like spring member mounted within said casing a-bove said upper end of said hammer, said spring member being at least twelve times as long at it is wide with its longitudinal axis substantially parallel to the longitudinal axis of said casing, said rod-like member having a lower end positioned to contact said upper surface of said hammer and receive energy from said hammer during a last part of the upward movement of said hammer and to return a portion of said energy during a rst part of the downrward movement of said hammer.

20. The percussive unit of claim 1'8 wherein the rotatable tubrular valve is connected to rotary drive means adapted to rotate said valve relative to the hammer.

21. The percussive unit of claim 1|8 wherein there is a flow restrictive means in said valve between the high pressure power uid inlet ports and the power fluid cxhaust ports, said flow restrictive means adapted to continuously pass a small amount of power fluid from an upper inlet section to a lower outlet section of said valve.

212. The percu-ssive unit of claim 118 wherein there is motion conversion means adapted to employ the reciprocating movement of said hammer to coordinate the axial position of said hammer with the circumferential rotative position of said ports in said valve during at least a portion of the reciprocating cycle of said hammer.

23. The peroussive unit of claim 22 wherein the motion conversion means is a cam and follower means, said cam having at least one curved surface positioned to contact said follower during at least a portion of the reciprocating cycle of said hammer.

24. The percussive unit of claim 18 wherein the central circular bore through the hammer has an inside diameter greater than the maximum width of the elliptical section of said valve to form a diametral clearance, and the difference between the major and minor axes of said elliptical surface being at least equal to said diametral clearance.

25. The percus'sive unit of claim 24 wherein the difference between the major and minor axes of said elliptical surface is at least twice the diametral clearance.

26. The percussive unit of claim 1.8 wherein the central circular bore through the hammer has an inside diameter greater than the maximlum width of the elliptical section of said valve to form a diametral clearance of at least 0.002 inch, and the difference between the major and minor axes of said elliptical surface is at least 0.002 inch.

27. The percussive unit of claim 26 wherein the difference between the major and mino-r axes of said elliptical surface is at least 0.004 inch.

28. In a power lluid-operated percussive unit of a rotary percussive drilling system having a hammer and a drilling means and wherein the percussive unit and the drilling means are rotated and the drilling means is irnpacted against the earth by axial cyclic reciprocation of the hammer, and the power fluid is passed through a tubular, rotatable valve having two high pressure power iluid inlet ports located in said valve in a position to alternately pass power iluid to passages communicating with the opposite ends of said hammer as said valve is rotated, and two power liuid exhaust ports located in said valve in a position to alternately vpass power uid from passages communicating with said opposite ends of said hammer to a power fluid exhaust passage as said valve is rotated, the improvement comprising said two high pressure power uid inlet ports being located on opposite sides of said valve and approximately 180 degrees on center so that said high pressure power fluid inlet ports are spaced substantially symmetrically around the surface of said valve, said two power fluid exhaust ports being located on opposite sides of saidl valve, said two power fluid exhaust ports being approximately 180 degrees on center from each other and approximately degrees on center from said high pressure power uid inlet ports, and a portion of the exterior surfaces of said valve between said high pressure power fluid inlet ports and said power fluid exhaust ports being in the shape of two eccentric arcs having offset centers of curvature, said two eccentric arcs being symmetrically spaced on opposite sides of said valve having an axis of symmetry the same as an axis of symmetry of said high pressure power uid inlet ports.

29. The improvement of claim 218 wherein the rotatable tubular valve is connected to rotary dn've means adapted to rotate said valve relative to the hammer.

30. The improvement of claim 28 wherein there is a ow restrictive means in said valve between the high pressure power uid inlet ports and power fluid exhaust ports, said ow restrictive means adapted to continuously pass a small amount of power tluid from an upper inlet section to a lower outlet section of said valve.

31. The improvement of claim 28 wherein there is motion conversion means adapted to employ the reciprocating movement of said hammer to coordinate the axial position of said hammer with the circumferential rotative position of said ports in said valve during at least a portion of the reciprocating cycle of said hammer.

32. The improvement of claim 3'1 wherein the motion conversion means is a cam and follower means, said cam having at least one curved surface positioned to contact said follower during at least a portion of the reciprocating cycle of said hammer.

33. The improvement of claim 218 wherein the valve passes through a central circular bore through the hammer, said central circular bore having an inside diameter greater than the maximum width of the ported section of said valve to form a diametral clearance, and the distance between the centers of curvature of the two eccentric arcs is at least twice said diametral clearance.

314. Improvement of claim 33 wherein the distance between the centers of curvature of the two eccentric arcs is at least four times the diametral clearance.

35. The improvement of claim Z8 wherein the valve passes through a central circular bore through the hammer, said central circular bore having an inside diameter greater than the maximum width of the ported section of said valve to form a diametral clearance of at least 0.002 inch and the distance between the centers of curvature of the two eccentric arcs is at least 0.004 inch.

36. The improve-ment of claim 35 wherein the distance between the centers of curvature of the two eccentric arcs is at least 0.008 inch.

References Cited by the Examiner UNITED STATES PATENTS 5/1959 Dulaney 173-73 9/ 1965 Spannhake 9l-40 

1. IN A POWER FLUID-OPERATED PERCUSSIVE UNIT OF A ROTARY PERCUSSIVE DRILLING SYSTEM HAVING A HAMMER AND A DRILLING MEANS AND WHEREIN THE PERCUSSIVE UNIT AND THE DRILLING MEANS ARE ROTATED AND THE DRILLING MEANS IS IMPACTED AGAINST THE EARTH BY AXIAL CYCLIC RECIPROCATION OF THE HAMMER, AND THE POWER FLUID IS PASSED THROUGH A TUBULAR, ROTATABLE VALVE HAVING TWO HIGH PRESSURE POWER FLUID INLET PORTS LOCATED IN SAID VALVE IN A POSITION TO ALTERNATELY PASS POWER FLUID TO PASSAGES COMMUNICATING WITH THE OPPOSITE ENDS OF SAID HAMMER AS SAID VALVE IS ROTATED, AND TWO POWER FLUID EXHAUST PORTS LOCATED IN SAID VALVE IN A POSITION TO ALTERNATELY PASS POWER FLUID FROM SAID PASSAGES COMMUNICATING WITH SAID OPPOSITE ENDS OF SAID HAMMER TO A POWER FLUID EXHAUST PASSAGE AS SAID VALVE IS ROTATED, THE IMPROVEMENT COMPRISING SAID TWO HIGH PRESSURE POWER FLUID INLET PORTS BEING LOCATED ON OPPOSITE SIDES OF SAID VALVE AND APPROXIMATELY 180 DEGREES ON CENTER SO THAT SAID HIGH PRESSURE POWER FLUID INLET PORTS 